The use of load compensation valves in braking systems for railway vehicles has been proposed in past installations. There is shown and disclosed one type of conventional load compensation valve in U.S. Pat. No. 4,586,754, issued on May 6, 1986, and assigned to the same assignee. The structure of the main portion of the load compensation valve is shown in FIG. 5. The load compensation valve 1 includes a supply chamber 2, an output chamber 3, a supply valve 4, an exhaust valve rod 5, and a check valve 30. This load compensation valve 1 may be used to control a brake system as is shown in FIG. 6. In practice, the amount of braking force or power exerted by the brake cylinders is dependent upon the load carried by the given car. As shown, the load compensation valve 1 includes the supply chamber 2 which is connected to an air pressure source MR via a three-way cock KV and also includes the output chamber 3 which is connected to a plurality of electromagnetic valves MV1, MV2 and MV3. A pair of control pressure lines AS1 and AS2 are connected to the lower end of the load compensation valve. The pressure level in each control line AS1 and AS2 is matched to the car load sustained by the air springs on the car. The pressure forces are transferred to the balance piston, which is not shown in FIG. 5, to act on the exhaust valve rod 5. If the control pressures AS1 and AS2 increase, the exhaust valve rod 5 is driven downwardly so that the exhaust valve rod 5 pushes on the supply valve 4 and opens the support port 11. Thus, the valve is placed in the supply position. When the supply pressure is fed into the output chamber 3 and when the pressure increases, the supply valve 4 moves toward its supply valve seat 10. When the air pressure in the output chamber 3 becomes substantially equal to the control pressures AS1 and AS2, the supply valve 4 sits on the seat and assumes a lapped position. In this lapped position, the control pressures AS1 and AS2 will decrease so that the exhaust valve rod 5 is driven in the direction in which it becomes separated from the supply valve 4. Thus, the exhaust valve seat 19 also separates from the supply valve 4. Accordingly, the air pressure in the output chamber 3 is exhausted through the exhaust port 18 to atmosphere. Thus, the air pressure decreases. Now, when the air pressure in the output chamber 3 becomes proportional to the control pressures AS1 and AS2, the exhaust valve rod 5 moves toward the supply valve 4 and results in a decrease in the air pressure in the output chamber 3. Again, the valve goes into the lapped position.
It will be seen that a check valve 30 is provided for causing a different action from the load compensation action. In the case where there is no check valve 30 and no first path 27 and no second path 28, it will be seen that during the time the load compensation valve is operating, the secondary pressure of the control pressure causes the supply chamber 2 to open to the atmosphere via the three-way cock KV. In this state, the supply valve 4 is maintained closed because of the structure of the load compensation valve 1. Thus, the pressure on the output side remains stable. Therefore, the check valve 30 is provided to overcome this condition. For example, in situations where the car has to be moved without a brake command, the brake is manually released so that the brake pressure is released. However, when it is manually released during the operation of the check valve 30, it may be dangerous when the pressure is operating on the output side.
The check valve 30 remains closed even though the supply valve 4 exhausts the supply chamber 2. This is because the check valve 30 is made in such a way that the air pressure from the source MR is operating on the supply valve 4 to balance it completely or almost completely. In order to improve the response, namely, the sensitivity of the check valve 30, it is necessary that it responds to movement of the control pressures AS1 and AS2. In other words, the difference between the effective area A1 of the supply valve seat 10 and the effective area A2 of the sliding part 4c of the supply valve 4 is made to be small so that the force of the air pressure of the source MR that is pushing on the supply valve 4 is small. Conversely, the sensitivity becomes satisfactory since the force required to open the supply valve 4 can also be small. When the air pressure of the source MR is P.sub.MR, the seating force F of the supply valve 4 can be calculated by the following equation: EQU F=P.sub.MR (A1-A2)
Therefore, in this case, the secondary pressure is operated by the control pressure, even though the air pressure in the supply chamber 2 is released by operation of the three-way cock KV. Accordingly, the required force to close the supply valve 4 does not depend on the air pressure of the source MR, so that it does not open.
If a check valve is provided, the pressure in the supply chamber 2 can be discharged by the operation of the three-way cock KV. Thus, the air pressure in the second chamber 24 can decrease so that the required force to open the valve is based on the difference of the air pressures between the first chamber 23 and the second chamber 24. The required force is larger than the force to close the check valve 30. Thus, the check valve 30 separates from the check valve seat 26 so that the passageway 25 is opened, and the air pressure in the output chamber 3 is conveyed through the first path 27, the first chamber 23, passageway 25, the second chamber 24, the second path 28, to the supply chamber 2. Since the air pressure in the output chamber 3 flows in the reverse direction, the air pressure in the output chamber 3 decreases. Accordingly, this action affects the balance piston so that the exhaust valve rod 5 moves in a direction to open the supply valve 4 while it is seated on the supply valve 4, and it opens the supply valve port 11 and this becomes the supply position, so that the secondary side pressure fluid is exhausted, in a reverse direction through the supply valve port 11.
Referring now to FIG. 6, it will be seen that the brake control mechanism BV, the supply train lines SB1, SB2 and SB3, the relay valve RV, the brake cylinder BC, the electromagnetic valves MV1, MV2 and MV3, and the electrical conductors and the pneumatic piping constitute the brake system. The control valve structure is substantially the same as the load compensation valve 1 of FIG. 5. The relay valve RV normally includes the supply chamber C4 and the output chamber C5.
In the above-mentioned load compensation valve 1 of the prior art, a problem occurs when dust settles on the seat surface of the check valve 30 since it cannot block-off the air flow. In addition, if the spring 31 becomes broken or is damaged the check valve 30 will not seat properly so that it cannot block-off the air flow. When the air flow cannot be positively blocked, the load compensation valve 1 will not work properly. In addition, the plate of the check valve 30 has a narrow guideway so that it sometimes becomes cocked during operation. In order to prevent this, the guideway should be enlarged to the size of the supply valve 4. However, this enlargement will result in the whole check valve 30 to become very large. Thus, the previous load compensation valve 1 and the related structural portions will substantially increase when the check valve 30 is incorporated in the load compensation valve so that the reliability of the valve system, as a whole, decreases the entire structure and becomes excessively large, which causes problems.